Hydraulic pump or motor



Nov. 7, 1961 T. BUDZICH 3,007,420

HYDRAULIC PUMP 0R MOTOR Filed 001;. 7, 1959 2 Sheets-Sheet 1 FIG.

INVENTOR. TADEUSZ BUDZICH Nov. 7, 1961 T. BUDZICH HYDRAULIC PUMP ORMOTOR 2 Sheets-Sheet 2 Filed Oct. 7, 1959 WSW m y Q 5 M M MW 8 UnitedStates Patent C) 3,007,420 HYDRAULIC PUMP R MOTOR Tadeusz Budzich, 3344Colwyn Road, Cleveland, Ohio Filed Oct. 7, 1959, Ser. No. 844,964 8Claims. (Cl. 103162) This invention relates to hydraulic apparatus andmore particularly to fluid pressure energy translating devices commonlyknown as fluid pumps and motors.

In still more particular aspect the invention relates to fluid pumps andmotors of the type having a cylinder barrel in which pistons arereciprocated as influenced by a cam plate device.

In still more particular aspects the invention relates to fluid pumpsand motors in which the pistons transmit the driving torque to or from acylinder barrel in the form of cantilever loads induced by transversecomponents of piston thrust when working against an inclined cam plate.

In a pump or motor of the above described types, transverse componentsof pistons thrust not only constitute the driving torque of the devicebut also result in setting up large transverse forces in the cylinderbarrel, tending to displace it as from its abutment with a valve plate.This disturbs the distribution of requisite sealing pressures (as on theflat face of such a valve Plate) reducing the efliciency of the unit andunder extreme conditions causing separation of cylinder barrel and valveplate.

In one arrangement of the prior art the cylinder barrel is supported bya radial bearing encircling it and transmitting the transverse forcesdirectly to the pump housing. This solution gives bulky and heavyconstruction of limited maximum speed as dictated by characteristics ofthe large bearing.

In another prior art arrangement the cylinder barrel is supported on ashaft retained by bearings in the housing. The point of contact betweenthe cylinder barrel and the shaft usually takes place at a spline, thelooseness in the spline providing limited universal freedom, In thissolution the transverse loads bearing against the spline teeth produce abinding effect and impair the efliciency of the universal action.

Here-tofore, too, in pumps and fluid motors, leakage along the pistonsis controlled by maintaining minimum clearances. But the tightclearances and long overlaps are synonymous with high costs, and alsoincrease the dimensions of the device. Such minimum clearances are alsodisadvantageous because easily damaged by dirt particles in the pumpedfluid and because of inability to carry away the heat generated at thepoints of high contact pressure. Another disadvantage of theconventional solution lies in the fact that due to cantilever loading ofthe piston, the piston assumes an eccentric position in relation to thecylinder bore. The leakage then, as is well known in the art, may be asmuch as three times greater than would be the case were the piston toremain concentric in the cylinder bore.

It is an object of the present invention to provide simple andinexpensive means for overcoming the above mentioned difliculties.

In accordance with one aspect of the present invention, transverseforces otherwise tending to upset the equilibrium of the cylinder barrelare balanced by proper selection of the point of its support along thelongitudinal axis of the pump more or less independent of the splinesand this feature cooperates with spring biased series leakage resistancemeans located in an opening or openings intermediate the cylinder boreends and around each piston, to reduce binding, reduce weight, reduceleakage, reduce costs, and greatly improve operational efliciency.

It is therefore another object of the invention to provide improvedmeans for supporting a cylinder barrel on a shaft in which bindingaction due to transverse loads is eliminated from the driving spline.

A further object of this invention is to reduce leakage of high pressurefluid along the pumping (or motoring) pistons. I

A still further object is to control leakagewhile increasing theclearance between piston and cylinder bore for economy.

Still another object of this invention is to provide separate readilyreplaceable means to reduce leakage about each piston, said means, ifdamaged being replaceable without replacement of the expensive cylinderbarrel or piston.

A further object of this invention is to mount sleeves providingresistance to leakage in such a way as to eliminate all the transverseloads acting on them so that under action of the reciprocating motion ofthe piston they will take a central position thereabout further reducingleakage.

Other objects and advantages will become apparent and the invention maybe better understood from consideration of the following descriptiontaken in connection with the accompanying drawing, in which:

FIG. 1 is a vertical cross-sectional view through an axial piston pump(or motor) provided with piston-leakage sleeves according to theinvention.

FIG. 2 is an enlarged view detailing the-construction of one suchleakage sleeve.

FIG. 3 is a graphical representation of force distribution in the planea-a of FIG. 2 on the sealing area be- Description Referring first toFIG. 1, I have shown a pump body comprising two sections, namely a pumphousing 10 and a pump cover 11, With the two suitably connected by bolts12. A valve plate 14 is secured to the pump cover 11 by dowel pins 15and has its outer periphery 16 engaging a bored recess 17 in the housing10. Valve plate 14 is provided with kidney shape arouate slots 19, 20(see FIG. 5) which are in direct communication with pump cover inlet andoutlet ports (not shown). If desired the valve plate may be providedwith leakage drillings 14a and the cover with communicating leakagepassages 11a.

A drive shaft 22 extends longitudinally through the pump and is retainedat the cover end in a bearing 23 and at the opposite end of the housingin a bearing 24.

Bearing 23 is retained in the cover by a shaft cover 25 having a sealingring 26. Bearing 23 is shown as a ball bearing which provides the endlocation for the shaft 22 retained with respect to'the hearing by twosnap rings 27 and 29. The other bearing (24) has its inner race againsta drive shaft shoulder 31 but may be of roller type (as shown) andprovide radial location and carry radial loads only. A reaction sleeve32 is tightly pressed on shaft 22 and against the shoulder 31. Sleeve 32has an enlarged end 34 spaced radially outward from the drive shaft 22to slidably engage a cylindrical extension portion 36 of a cylinderbarrel 37,

The cylinder barrel 37 is provided with cylinder bores 38 and pistons 39work in operational contact with these bores. Each piston 39 terminatesin a spherical end 41 and a plurality of shoes 42 closes each over oneof the spherical ends 41.

As shown, the shaft 22 has a male spline portion 44 engaging acorresponding spline located in the extension 36 of cylinder barrel 37.It will be seen that cylinder barrel 37 is radially spaced from shaft 22except in the region of the splines. The longitudinal bores 38 areinterrupted each by the same circular recess indicated generally at 46and for housing piston return springs 47 and leakage sleeves 48. At theright end, as viewed in FIG. 1, the cylinder bores 38 communicate withrecessed reliefs 49 terminating in slots 50 which pass through theendmost flat face of the cylinder barrel 37 (see FIG. 4). This flat facealso has usual dynamic pads 51 and leakage slots 52 and works inoperational contact with a flat face of the valve plate 14.

Within the pistons the various drillings such as 54 and 55 connect thecylinder bores 38 with an opening 56 in the respective piston shoe 42and with the balancing area 57 thereof. The piston shoes 42 with thebalancing areas 57 and balancing lands 58 work in sliding engagementwith the angled flat face of a cam plate 60 radially located at one endof the housing by a flange portion 61 engaging a corresponding recess inthe housing 10.

The piston return springs 47 are anchored at one end each to therespective piston 39 by a snap ring 63 and a spring retainer 64 (seeFIG. 2). The other end of each piston return spring 47 engages a flangeportion 65 of the corresponding leakage sleeve 48. The leakage sleeve 48encircles the piston 39 and, in cooperation with the spring retainer 64,radially locates the piston return spring 47 while the return spring 47keeps the leakage sleeve 48 in abutment with the cylinder barrel 37.With this construction, and as hereafter explained, any leakage oil,flowing along a clearance 66 (between piston 39 and cylinder bore 38 asindicated in FIG. 2) must also flow through a clearance 67 between theleakage sleeve 48 and the piston 39. Since resistance to leakage flow isproportional to length of flow path, the leakage sleeve is purposelyconfigured to extend a material distance longitudinally to reduceleakage volume.

Operation Although the device may function either as a pump or as amotor, its operation will be described as that of a pump. Accordinglythe drive shaft 22 may be assumed connected to a suitable prime mover,not shown, and the inlet and outlet ports of the pump connected to ahydraulic system. The drive shaft 22, revolving in the bearings 23 and24 will induce rotation of cylinder barrel 37 through the male andfemale splines. The cylinder barrel, while revolving, will sequentiallyregister each cylinder bore with inlet and discharge ports of the pump.In a well known manner the cam plate 60 will cause the pistons toreciprocate in timed relation with this sequential registration andthereby will produce a pumping action. The proper sealing between themating faces of the cylinder barrel 37 and valve plate 14 is affected byhydraulic pressure, acting within cylinder bores 38, and force due tothe piston return springs 47.

The floating cylinder barrel 37 is free to slide in a longitudinaldirection and align itself to the valve plate. The longitudinal forces,acting on it, will produce an effect of rapidly closing any gap existingbetween these mating members. The flat face of cylinder barrel 37,mating with valve plate 14, is hydrostatically balanced in a well knownmanner so that the actual contact pressure, existing between these twoparts, is caused by only a small fraction of the hydraulic reactionforce available.

The cylinder barrel circular recess 46 defines a central tubular part 68connecting a cylinder barrel front flange part 69 with a rear parthousing the pressure bores. While each of the circumferentially spacedcylinder bores 38 extend from thefront flange 69 to the main body of thecylinder barrel, each is interrupted by the circular recess 46 housingpiston return springs 47.

As already mentioned longitudinal passages such as 54 and 55 extendthrough each piston and connect with a passage 56 and balancing area 57of the respective piston shoe 42 to affect in a well known manner ahydrostatic balance. The distribution of hydraulic forces, acting on theflat faces of piston shoes 42 is so arranged that during the dischargestroke there exists a small resultant force keeping the piston assemblyagainst the fiat face of the cam plate 60. During the suction stroke,the piston assembly is maintained in contact with the cam plate 60 bythe piston return springs 47 working in abutment with the cylinderbarrel 37 through the leakage sleeve 48. For maximum pumping efliciencythe cylinder barrel 37 must be free to align itself against thecorresponding face of the valve plate 14, irrespective of deflection inthe housing 10 and out-of-squareness of sealing surfaces due tomanufacturing tolerances.

Referring to FIG. 1, a force FF is induced in the piston assembly bypressure generated in the cylinder bore 21, during the discharge stroke,acting on the piston area. This force is opposed by a force FC,perpendicular to the surface of cam plate 60 and a force FR, actingthrough the center of spherical surface 41, and transverse to the axisof the piston. The forces FR, acting on all pistons subjected to thedischarge pressure, are transmitted in the form of cantilever loads tothe cylinder barrel extension 36 and to the cylinder bores 38. Theseforces produce the driven or driving torque depending on whether thedevice is acting as a pump or as a motor. At the same time the forces FRproduce a transverse moment acting on the cylinder barrel, tending tounseat it from its abutment with the valve plate 14. In known mannerthese moments can be eliminated from the cylinder barrel by arrangingthe point of support between cylinder barrel and drive shaft at thepoint of intersection of a plane connecting the centers of the sphericalpiston ends with the center line of the cylinder barrel. Cylinder barrelcylindrical extension 36, with its flat land 69, engages in a slidablemanner the inner periphery of the enlarged portion 34 of the reactionsleeve 32. The sum of the forces FR, represented by a resultant forcesFRR, see FIG. 1, is transmitted from the cylinder barrel to the innersurface of the reaction sleeve 32 and then to drive shaft 22. The centerof contact pressure between cylinder barrel 37 and reaction sleeve 32 isshown at point 70. The splined connection between the cylinder barrel 37and shaft 22 is located directly inside the reaction sleeve 34, thelength of this splined engagement being equally spaced on each side ofthe center of contact pressure represented by point 70.

In a conventional type of solution the cylinder barrel is supporteddirectly on the shaft in the region of the splines without any suchreaction sleeve to assist. This type of cylinder barrel suspensionsuffers from very serious disadvantages. Large transverse loads, actingon the involute profile of the spline teeth, tend to produce a bindingaction, which very seriously affects the freedom of alignment of thecylinder barrel. At the same time these transverse loads produce aneccentric location of the female spline in relation to the male splineby the amount of clearance available in the drive. Since the load issupported on the full length of the spline teeth any adjustment in theposition of the cylinder barrel, in relation to its longitudinal axis,will not only result in a binding action but it will induce anadditional couple disturbing the equilibrium of the cylinder barrel. Thepresent invention eliminates this binding action on the spline byremoving from the splines the transverse loads acting on the cylinderbarrel, as transmitted from the cantilever piston assemblies. The greatcircle engagement of the cylindrical extension 34 of the reaction sleeve32 locates the cylinder barrel 37 in relation to the drive shaft 22 andabsorbs the resultant force FRR. The reaction sleeve 32 ensures that thefemale spline is kept concentric with the male spline and this resultsin the spline teeth carrying torque loads only. This gives a muchgreater freedom of alignment of the cylinder barrel. The land 69provided on the cylindrical extension 36 of the cylinder barrel 37 ismade to engage the reaction sleeve 32 directly at the mid-lengthposition of the spline. Only in this type of cylinder barrel suspensioncan the maximum of movement of cylinder barrel be obtained for anyspecified clearance in the spline teeth.

Thus the transverse forces upsetting the equilibrium of the cylinderbarrel are balanced by proper selection of the point of its supportalong the longitudinal axis of the pump. The point of support of thecylinder barrel is substantially at the point of intersection of itslongitudinal axis with the plane connecting the centers of the sphericalpiston ends so that the cylinder barrel mounted around this point willbe capable of aligning itself against the flat surface of the valveplate irrespective of deflection vof the housing and out-of-squarenessof sealing surfaces due to manufacturing tolerances, but unlike priorart arrangements in the described arrangement this does not involve thesplines.

To reduce leakage and therefore increase efliciency of a piston pump,tight clearances between the pistons and cylinder bores are usuallymaintained. 'Ihe tight clearances'of the prior art, although beneficialfrom the standpoint of performance, carry certain inherentdisadvantages. They are expensive to produce and susceptible to damageby dirt particles contained in the fluid pumped. Tightly fitted pressurebores must not only seal but also carry the side loads to which thepistons are subjected. In cylinder bores working with tight clearancesit is difficult to cool the cylinder walls subjected to high localizedpressures. A strict compromise between efliciency and life of thesealing surface becomes necessary. There is one additional disadvantagecommon to all piston pumps in which the pistons are carrying transverseloads. Under the influence of these loads the pistons assume aneccentric position in the clearance provided in the cylinder bore. It iswell known in the art that the leakage past these eccentrically locatedpistons will be as much as three times larger than would be the casewere these pistons to remain concentric in the cylinder bores. Thepresent invention eliminates all these disadvantages, as listed above,by providing additional resistance to leakage in series with theresistance to flow in the clearance between each piston and its cylinderbore in form of a leakage sleeve 48. Each leakage sleeve 48 encirclesthe respective piston and under action of the piston return spring 47,is kept in abutment with the flat face of the cylinder barrel 37. Theactual decrease in the leakage volume will be proportional to the lengthof the leakage sleeve, length of the piston overlap in the cylinderbore, and respective clearances in leakage and in cylinder bore. Becauseof its construction the leakage sleeve can be more cheaply manufacturedto closer tolerances and therefore the clearance between this leakagesleeve and the piston can be kept to a minimum. At the same time theclearance between the piston 39 and the cylinder bore 22 can beincreased, thus reducing the cost of a large and complicated part. Thisincrease in clearance between the piston and the cylinder bore, with theuse of leakage sleeves, does not decrease the efficiency of the pump andat the same time it provides a more favorable condition of cooling andloading of the rubbing surfaces. An important advantage of this solutionis that (although, under action of the catilever transverse load, thepiston will take an eccentric position in the cylinder bore) the leakagesleeve can remain concentric, further reducing the total leakage alongthe piston. The ability of the leakage sleeve 48 to maintain its centralposition in respect to pump piston is obtained by slideably mounting asleeve sealing surface 72 along the cylinder barrel face AA (see FIG.2).

The leakage oil when forced to flow by the pressure gradient through theclearances 66 and 67 (FIG. 2) along the surface of the piston 39 affectsthe pressure distribution on sealing surface 72 at the point of contactbetween leakage sleeve 48 and cylinder barrel 37. The pressuredistribution on sealing surface 72 is shown as P on the right of line AAin FIG. 3. The pressure P acting on the inner edge of the sealing areais equal to discharge pressure less than pressure drop in clearance 66.The resultant force FL caused by the pressure distribution on sealingface 72 tends to separate leakage sleeve 48 from the face AA of cylinderbarrel 37. This separation in prevented by a force FS representing theminimum preload in the piston return spring 47. The force distributionon the sealing face 20 is so arranged that at all times force FS islarger than force FL resulting in effective sealing. This constructionalso permits replacement of individual leakage sleeves, if damaged,making wear in the cylinder bores 38 of the cylinder barrel 37 lessimportant. At the same time this solution permits a wider use ofmaterials in construction of the cylinder barrel and leakage sleeves.

There is thus provided a device of the character described capable ofmeeting the objects above set forth and whereby improved mounting ofcylinder barrel and additional spring biased resistance means to preventleakage (despite large tolerances between each piston and piston bore)as well as the large openings between portions of each cylinder borecooperate to minimize or obviate binding action due to transverse forcecouples, to reduce weight of rotating parts, to minimize leakage of highpressure fluid and thus to increase efliciency of the device, and toafford substantial economies both in manufacture and in maintenance.

While I have illustrated and described a particular embodiment, variousmodifications may obviously be made without departing from the truespirit and scope of the invention which I intend to have defined only bythe appended claims taken with all reasonable equivalents.

I claim:

1. In an energy translating fluid pressure device comprising a housing,a drive shaft journalled with respect to the housing, a cylinder barrelrotatable with said shaft and having a plurality of cylinder bores andpistons reciprocable therein, valve structure having inlet and dischargeports which sequentially register with each cylinder bore as thecylinder barrel rotates, spherical piston ends associated with thepistons outside the cylinder barrel, shoes closing one over each of thespherical piston ends while universally mounted with respect thereto,and a cam plate operably connected with the shoes and mounted withrespect to the housing, the combination of said barrel and said driveshaft having drivingly interconnected splines, said shaft includingconstraining means, said constraining means being in axially-slidableengagement with said barrel at a region radially spaced from said splineinterconnection, whereby said constraining means permits axial movementbetween said shaft and said barrel but limits transverse movementtherebetween.

2. In a device as in claim 1, the combination thereof furthercharacterized by the constraining being a sleeve having an inner bore atone end engaging the shaft and an enlarged opposite end with an innerbore engaging a portion of the cylinder barrel radially spacing saidcylinder barrel from the shaft.

3. In a device as in claim 1, the combination thereof furthercharacterized by the cylinder barrel having a cylindriral extensionhaving inner peripheral female splines for mating with male splinesprovided on the shaft, while said cylinder barrel extension has an outerperipheral land, and constraining means, said constraining meanscomprising a sleeve engaging the shaft at a region axially spaced fromthe male splines at one end and with its other end slideably engagingsaid peripheral land of said cylinder barrel.

4. A pressure fluid mechanism comprising a housing, a drive shaftjournalled in said housing, a rotatable cylinder barrel having aplurality of axially extending cylinder bores each interrupted and madeinto fore and aft support and pressure portions by a circular recessextending around the cylinder barrel, pistons mounted for sliding insaid cylinder bores and each having a partspherical surface at one end,piston shoes universally mounted one on each of said part-sphericalpiston ends, a cam plate having an inclined face arranged to engage theshoes and thus to move the pistons in at least one direction withrelative rotation of cylinder barrel with respect to the cam plate,resilient biasing means located in the barrel circular recess andresiliently biasing each piston with respect to the cylinder barrel foraiding motion of the pistons in one direction, a relatively stationaryvalve plate arranged in abutment with an end of the cylinder barrel andhaving inlet and outlet ports for sequentially registering with eachcylinder bore as the barrel rotates, cylinder barrel rotation meanscomprising a male spline on the drive shaft and a co-operating femalespline on the cylinder barrel, and a sleeve mounted for rotation withthe drive shaft and slidably engagnig a portion of the cylinder barrelcontaining the female splines and radially spaced therefrom with themidpoint of engagement coinciding with a plane passing through themidpoint of the female splines transverse to the axis of the shaft.

5. A pressure fluid mechanism as in claim 4 further characterized by asubstantial clearance provided between male and female splines, thecylinder barrel outside the splines being radially spaced from the shaftto ensure requisite universal action between cylinder barrel and driveshaft, while the constraining means maintains a mid-point of the splinesconcentric to obviate binding action under transverse load.

6. A pressure fluid mechanism as in claim 4 further characterized by aplurality of sleeves each encircling a dilferent piston and located inthe circular recess adjacent the cylinder barrel pressure portion, saidsleeves and barrel having mating pressure sealing faces, the'sealingface of each sleeve having an inner diameter less than the diameter ofits associated cylinder bore, and the biasing means comprising aplurality of springs one about each of said sleeves and against aportion thereof to urge said mating faces into engagement while theopposite end of each spring is secured with respect to the associatepiston, with the area of said faces and the preload force of saidsprings so selected that the force due to pressure gradient acting onsaid faces is less than the preload in said springs at any pistonposition.

7. In the combination of claim 1, the centers of the spherical pistonends being co-planar in all operative positions and wherein the plane ofthe centers of the spherical piston ends intersects axial slidingengagement of the constraining means and the cylinder barrel and alsointersects the driving interconnection of the splines of the barrel andthe shaft in all operative positions of the device.

8. In an energy translating fluid pressure device comprising a housing,a drive shaft journaled with respect to the housing, a cylinder barrelrotatable With the shaft and having a plurality of cylinder bores andpistons reciprocable therein, valve structure having inlet and dischargeports which sequentially register with each cylinder bore as thecylinder barrel rotates, spherical piston ends associated with thepistons outside the cylinder barrel, shoes closing one over each of thespherical piston ends while universally mounted with respect thereto,and a cam plate operably connected with the shoes and mounted in thehousing, the combination of the barrel being disposed around at least aportion of the shaft, the shaft and the barrel each having splinesdrivingly interconnected, said shaft having an annular constrainingsleeve, said sleeve ahving an inner cylindrical surface radiaHy spacedfrom the shaft spline and defining an annular recess between the shaftspline and the sleeve, said barrel having an annular'projecting portion,the splines of the barrel being located on the inner rim of saidprojecting portion, the outer rim of said projecting portion being anannular surface mated to coact with the inner surface of said sleeve inaxially sliding engagement, said projecting portion projecting into andterminat' ing in said recess, whereby axial movement between the barreland the shaft is permitted but transverse movement therebetween islimited.

References Cited in the file of this patent UNITED STATES PATENTS2,299,235 Snadel et al. Oct. 20, 1942 2,817,954 Badalini Dec. 31, 19572,845,876 Keel Aug. 5, 1958 2,896,546 Lundgren et al July 28, 1959

